Rotary gear motor/pump having hydrostatic bearing means

ABSTRACT

A gear hydraulic pump or motor utilizes a sun gear, ring gear, and pairs of planet gears in a planetary arrangement with power shafting connected to either sun or ring gear or both. All gears are centered in their clearance by hydrostatic bearings supplied with fluid from the tooth trapping zones.

United States Patent 1 Rumsey [451 Sept. 3, 1974 1 i ROTARY GEAR MOTOR/PUMP HAVING HYDROSTATIC BEARING MEANS [76] Inventor: Rolling Douglas Rumsey, 148

Summer St., Buffalo, NY. 14222 [22] Filed: Jan. 2, 1973 21 Appl. No.: 319,978

Related US. Application Data [63] Continuation-in-part of Ser. No. 120,890, March 4,

1971, abandoned.

[52] U.S. Cl 418/77, 418/79, 418/81, 418/189 [51] Int. Cl...... F0lc 21/00, F030 3/00, F04c 15/00 [58] Field of Search 418/77-79, 418/81, 189

[56] References Cited UNITED STATES PATENTS 2/1958 Toyoda ..418/79 92/1958 Wakeman ..418/77 5/1959 Haberland 418/79 2,997,960 8/1961 Kimijima et a1. 418/79 3,123,012 3/1964 Gilreath 418/189 3,170,408 2/1965 Hill et al. 418/77 3,447,472 6/1969 Hodges et a1. 418/77 FOREIGN PATENTS OR APPLICATIONS 739,357 10/1955 Great Britain 418/79 Primary Examiner-William L. Freeh Assistant ExaminerJohn J. Vrablik Attorney, Agent, or FirmChristel & Bean [5 7 ABSTRACT A gear hydraulic pump or motor utilizes a sun gear, ring gear, and pairs of planet gears in a planetary arrangement with power shafting connected to either sun or ring gear or both. All gears are centered in their clearance by hydrostatic bearings supplied with fluid from the tooth trapping zones.

7 Claims, 12 Drawing Figures PATENTED 35? 31974 SEEK-2N 3 ROTARY GEAR MOTOR/PUMP HAVING HYDROSTATIC BEARING MEANS CROSS REFERENCE TO RELATED APPLICATION This application is a continuation-in-part of my application Ser. No. 120,890 filed Mar. 4, 1971 for Rotary Gear Motor/Pump, now abandoned.

BACKGROUND OF THE INVENTION 1. Field of the Invention This invention relates to the field of fixed displacement rotary pumps and motors primarily adapted for usage in the general fluid power industry.

2. Description of the Prior Art Toothed wheels have been used as pumping devices I torque may be transmitted at either of two speeds defor many years and the major problems relating to their usage have been internal leakage leading to low volumetric capacity, trapping of fluid between the teeth resulting in shock forces resulting in rapid bearing wear as well as tooth cavitation, and unbalanced hydraulic forces on the gears which necessitated very high capacity bearings.

In order to overcome the above mentioned problems various improvements have been utilized in the past, including (to reduce leakage) pressure loading wear plates against the gear side faces which although effective in pumps results in such high starting friction in motors that their use is limited. Other techniques to reduce leakage include the employment of soft alloy housings into which the gears may cut their way to a perfect fit and the utilization of a floating internal gear pressured against a driving pinion by a pressure loaded bearing shell and crescent divider.

A great many inventors have attacked the problem of trapping of fluid between gear teeth and satisfactory solutions have been found in the form of T or t shaped notches cut into the side plates in the tooth mesh area which act to vent the trapping area to the inlet or outlet side of the pump. Helical and heringbone gears are also used to effect a solution.

Nearly all commercial rotary gear pumps utilize only two gears because of the increased manufacturing complexity involved when more are employed; nevertheless a number of units do employ planetary layouts for various purposes. Typical are the arrangements of US. Pat. Nos. 2,934,044 and 3,057,161 where multiple speeds are achieved by differentially valving the various pumping elements. A somewhat similar arrangement is employed in a hydraulic slip clutch.

SUMMARY OF THE INVENTION This invention relates to gear type hydraulic pump or motor mechanisms and is particularly concerned with such units having a planetary arrangement consisting of sun, ring, and planet gears; the particular embodiment being most effective from a porting standpoint when even pairs of planet gears are employed.

The present invention is primarily characterized by its high power density, compared to conventional gear type hydraulic machines. Inasmuch as each gear mesh constitutes a pumping action, the invention with for example four planet gears plus sun and ring gear achieves eight meshing areas, hence has eight imes the pumping capacity of a two gear unit or four times the capacity per gear. Similarly a two planet unit would have twice pending upon whether the sun or ring gear are engaged.

Other objects, features and advantages of the present invention will be readily apparent from the following detailed description of certain preferred embodiments thereof taken in conjunction with the accompanying drawings in which:

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a longitudinal developed sectional detail view through a gear pump/motor embodying features of the invention and taken along substantially line 1 1 of FIG. 2.

FIG. 2 is an end elevational and sectional detail view of the motor taken substantially along line 11 ll of FIG. 1.

FIG. 3 is an end elevational and half sectional detail view taken along line 111 111 of FIG. 1;

FIG. 4 is an enlarged fragmentary sectional detail view of a gear meshing area, taken about on line IV IV of FIG. 5;

FIG. 5 is a fragmentary sectional detail view taken substantially on line V V of FIG. 4;

FIG. 6 is an enlarged fragmentary sectional view like that of FIG. 4 but showing a modification;

FIG. 7 is an enlarged fragmentary sectional view taken about on line VII VII of FIG. 6;

FIG. 8 is a longitudinal developed sectional view through an alternate embodiment of the invention similar to FIG. 1 but incorporating power shafts attached to both sun and ring gears;

FIG. 9 is an enlarged fragmentary sectional view showing a detail of FIG. 7;

FIG. 10 is a fragmentary sectional view, similar to that of FIG. 5, but extended;

FIG. 11 is a fragmentary detail sectional view like that of FIG. 9 but showing a modification; and

FIG. 12 is a fragmentary elevational view of the check valve of FIG. 11.

DETAILED DESCRIPTION OF THE ILLUSTRATED EMBODIMENTS A rotary gear hydraulic motor/pump embodying features of the invention and as exemplified in FIGS. 1 and 2 has a housing comprising a cup shaped body 10 closed at its open end by a manifold plate closure 11 to which is bolted a second manifold 12 and between which are located various gear pumping elements consisting of an internal ring gear 13, an even number of planet gears 14, 15, 16, 17 meshing therewith, a sun gear 18 engaging all planets, sealing and spacer blocks 19, 20 filling most of the void between gears, hollow shafts 21 on which said planet gears rotate and wear plates 22 and 23 between which the gears are confined.

Flange member 24 which should be thin and/or spoked to isolate forces other than torque connects ring gear 13 to output (input-pump) shaft 25, supported in bearings 26 and 27.

Most gear hydraulic machines are fully reversible hence in a pump reversing the rotation reverses the flow and conversely in a motor reversing the flow switches the rotation, therefore there is no discharge and suction, per se. In manifold closure 11 there are two fluid ports 28 and 29 and for purposes of description 28 will be called the inlet or pressure port and 29 the outlet or return port. In a gear machine, the inlet in relation to the mesh is on that side at which the teeth are separating. On a gear machine as shown in FIG. 1 and FIG. 2 there are eight inlets and eight outlets since the unit consists of eight separate pumping meshes, and although circuitry other than parallel could be employed with the concept shown, the discussion will be confined to parallel operation.

Since each mesh constitutes a pump, there are four with the ring gear and four with the sun gear; hence four inlets and outlets must be piped to the outer meshes and an equivalent number to the inner.

Fluid entering at inlet 28 is conducted through manifold passages 30 (refer also FIG. 3) to gear mesh zone 31, where it divides with half proceeding across the teeth to passage 32 in manifold and hence to gear inlet 33. As the gears rotate under fluid pressure sun and ring clockwise in FIG. 2 fluid entrained between the teeth is carried by the sun gear from zone 31 to 34 where it exits into passage 35 in manifold 12 and reenters exit zone 36 where it rejoins the fluid carried by planet gear 16s teeth and together exit into passage 37 in manifold 11 thence proceed to outlet 29. Similar action occurs at the other gear porting areas.

In most fluid power devices where low friction and long operating life of relatively moving parts is a requirement, clearance sealing between the high and low pressure zones is normally used. Since it is desireable to maintain leakage losses through such clearances at a minimum, it is standard practice to keep diametral clearances below .002 inches and frequently in the low ten-thousands of an inch range, particularly where pressures above 1000 psi. are employed. As a result of the above, clearance leakage flow is nearly always laminar and as such varies as a function of the cube of the clearance.

Thus in hydraulic machinery such as gear pumps and motors, the leakage past the side of the gears will be seven times as great if the gears are touching one side of their housing and leaving all the leakage clearance on the other side than if the gears were centered in the clearance. It is a perculiarity of mechanisms to always operate in the mode of least work, therefore the gears will always operate in the first mode mentioned above unless forced to do otherwise. It is one of the features of this invention to provide a simplified means for centering the gears in their clearance, thereby reducing clearance leakage and overcoming frictional losses and the resultant wear attributable thereto.

In gear mechanisms using accurately manufactured spur gearing there are always more than two teeth in contact in the mesh zone, therefore when such gears are employed in pumps or motors fluid becomes trapped between these teeth. As the teeth come in to mesh, the action is to compress the fluid until a point on a line joining the gear centers is reached, beyond which expansion occurs. On well built equipment trapping pressures of many thousands of pounds per square inch can be generated and since the machine is not designed to withstand it, it normally will leak off, due to shaft and bearing deflection, bulging of the case etc., all of which causes rapid wear because of the intermitten peak overloads. As the teeth begin to separate after the high pressurized fluid has leaked away, cavitation immediately occurs resulting in cavitation erosion of the gear teeth and adjacent wear plates. It is thus seen that in each mesh, first a fluid compression cycle occurs immediately followed by a suction or expansion cycle. Normal gear pump practice is to connect the above mentioned compression zone to the expansion zone by means of small passageways in the side plates thereby tending to offset one with the other. Obviously the results are imperfect because the rates of compression and expansion are changing continually and oppositely.

Referring to FIG. 4, if the gear teeth are moving in the direction of the arrows, pocket 43 is closing, pocket 44 is enlarging and pocket 45 being on center has just stopped closing and will begin to enlarge. It will be noted that three teeth on gear 14 are in contact with two on gear 18 and that there are four points of contact, therefore the three above mentioned pockets are all isolated from one another. It is thus recognizable that two separate pumping areas occur, one near the tooth roots of gear 14 and the other at the tooth roots of gear 18. If openings 46 and 47 are provided in the side walls provided by plates 22 and 23, in communication with pockets 43 and 45, respectively, the fluid compressed or pumped in pocket 43 will be pumped through restricted passages 48 in plates 22 and 23 into annular hydrostatic bearing channels 49 in opposite sides of gear 14, thereby relieving trapping pressures and hydrostatically centering gear 14.

The hydrostatic bearings provided by channels 49 on opposite sides of each planetary gear 14 are in communication through the associated restricted passages 48, collector openings 46 and trapping pocket 43. If the gear 14 is off center, touching member 22, for example, a relatively large leakage path is provided between gear 14 and member 23, of much greater cross-sectional area than the feed passage 48 on the touching side, resulting in a pressure unbalance between the opposite sides of the gear causing the latter to shift on its axis into a centered position of pressure quilibrium. This, the gears 14 are centered between members 22 and 23 hydrostatically by the pressure fluid bearing recesses 48 on opposite sides of the gear.

A similar function is performed by pocket 45 which communicates with openings 47 and feeds annular hydrostatic bearing channels 51 in opposite sides of gear 18 throughrestricted passages 50.

Cavitation is avoided by providing vacuum relief recesses 52 in plates 22 and 23 which connect the vacuum trapping pockets 44 with any convenient source of hydraulic fluid for example the zone immedialy following the pocket 44 whereby fluid is freely supplied to the expanding pocket 44 in a manner avoiding the problem of cavitation.

It is readily apparent from the above that each gear mesh provides two trapping area pumping functions. With the six gears shown in FIG. 2 having eight meshing areas, there are sixteen pumping zones and in order to center the six gears, twelve hydrostatic bearing annuli are required. It is a simple matter to supply all twelve, preferably by providing dual feeds to hydrostatic bearing annuli in the sun and ring gears.

For example, it will be seen from FIG. that each hydrostatic bearing annulus 49 of each planetary gear 14 will be supplied from at least one pocket 43 via an opening 46 and a restricted passage 48. Both sides of each planetary gear can be supplied from each pocket 43, as in FIG. 5, but that is not necessary. Similarly, each hydrostatic bearing annulus 51 in sun gear 18 will be supplied from at least one pocket 45 via an opening 47 and restricted passage 50. Here again, both sides can be supplied from each pocket, but that is not necessary and alternating pockets can supply the opposite sides of the sun gear as illustrated in FIG. 10 in connection with the planetary gear 14 shown therein.

Ring gear 13 also is provided with channels 54 in the opposite sides thereof, providing annular hydrostatic bearings which are supplied with hydraulic fluid from a pocket 45' corresponding to pocket 45 but found between each planetary gear 14 and ring gear 13 where they mesh, whereby in the arrangement illustrated in FIG. 2, there will be four such pockets. Fluid trapped in pockets 45 is delivered via collector openings 47 and restricted passages 50' in members 22 and 23 to the hydraulic bearing annulii 54. Here again, each pocket 45 can supply both bearing channels 54, or the latter can be supplied from alternating pockets 45 as illustrated in FIG. 10 in connection with the planetary gear 14 shown therein.

It can thus be seen that each side of each gear in its hydrostatic bearing annulus is supplied continuously with a source of nearly unlimited pressure although of relatively low volume, and that if the gear tends to remain against the side plate it will be forced away, while on the opposite side where the clearance is large the small flow will leak off to the low pressure side of the unit, and to the hub and shaft bearing areas, without building up appreciable pressure, hence a centering action is assured.

The same procedure as outlined above can be employed on a standard two gear pump or motor, however since there are only two trapping areas pumping functions and four hydrostatic bearing annuli it becomes necessary to split the flow from each pumping area by means of a downstream orifice or flow control valve means.

The arrangement of FIG. 4 is unidirectional, and can be used with a two gear pump or motor, or in a larger unit. FIG. 6 illustrates a similar arrangement for a bidirectional pump or motor, which also can be two gear or larger. A planetary gear 60 is mounted in engagement with sun gear 61, both of which are mounted for rotation between wear plate members 81 and 82, in the same manner as gears 14 and 18 between members 22 and 23. Gears 60 and 61 are provided with annular hydrostatic bearing recesses 49' and 51' on opposite sides thereof, as shown in FIG. 7, and these are provided with fluid under pressure being pumped from the trapping pockets, as in FIG. 4, but through openings 72 and 73 and restricted passages 74 and 75, respectively, in plate 81. Similar openings and recesses are provided in plate 82, and assuming rotation in the direction of the arrows, the centering operation is the same as previously described. If the direction of rotation is reversed, the pressure balancing and centering function is accomplished by corresponding openings 90 and 91 and restricted passages 93, 94 on the opposite sides of the vertical center lines, in a manner readily understood upon assuming rotation of gears 60, 61 in the direction opposite to that indicated by the arrows in FIG. 6.

In this embodiment cavitation is avoided by the provision of a vacuum relief arrangement illustrated in FIGS. 6 and 7, and in FIG. 9 which show a port 83 through wear plate 81, placing the opening 72 in communication with an enlarged opening 84 in the opposite side of plate 81. A check valve in the form of a pan shaped disk 85 is held in place in opening 44 by a seal ring 86 which resiliently urges disk 85 against a housing portion 79. A channel 87 terminating in an annular groove 88 around a central hole or aperture 89 in disk 85 communicates with the drain side of the unit for example the shaft bearing clearance area which is a source of hydraulic fluid.

Disk 85 is formed of thin sheet material, such as steel of approximately .003 inch thickness which normally resiliently engages the seating surface of housing portion 79 surrounded by the groove 88, thereby closing off communications between channel 87 and port 83. Such a check valve arrangement with associated channel and port is provided for each trapped zone opening 72, 73, 90 and 91. Those check valves 85 which communicate with the trapping pockets remain closed, because their self closing action is reinforced by the pressure of the trapped fluid being pumped to the hydrostatic bearing recess. Those check valves 85 which are located on the opposite side of the vertical center line and which communicate with the expanding pockets, for example the check valve associated with opening 90 in FIG. 6, assuming rotation in the direction of the arrows, will open, however, because their self closing resiliency will be overcome by the reduced pressure in the expanding pocket, drawing fluid from the bearing, drain area through passage 87, past disk 85 which oilcans away from the housing seating surface, permitting fluid to flow through opening 89 and passage 83 to the expanding pocket, to avoid cavitation.

FIGS. 11 and 12 illustrate a modified check valve arrangement, wherein the valve 54 is in the form of a spider spring disk held in its recess by a chamferd retainer ring 55 in an opening such as 47 which can be enlarged. The restricted passage to the bearing channel can be in the form of drilled openings 50 communicating with opening 47 and with the bearing channel. A channel 87 from a source of fluid such as a drain or bearing area communicates with a passage 83 normally closed by the resilient engagement of disk 54 against the seat on member 23 surrounding passage 83, which seat is in turn surrounded by an annular passage communicating with the restricted passage 50' and with opening 47 through the semicircular cut out portion of the spider disk.

When the spring disk 54 is exposed to the reduced pressure in an expanding pocket it will open, permitting the delivery of fluid to the pocket to avoid cavitation. It will be appreciated that the passages 87 leading to the fluid source are substantially larger than the restricted delivery passages such as 50, 50' leading to the hydrostatic bearing recesses, whereby the hydrostatic bearing pressure needed for centering will be maintained as fluid is drawn from the source to avoid cavitation.

An interconnecting channel can be provided, as shown at 92 in FIG. 6, between the trapped fluid collecting openings, whereby one trapped zone or pocket can supply the hydrostatic bearing recesses on opposite sides of both meshing gears, in which case only a single check valve per side would be required.

FIG. 8 illustrates a modified embodiment incorporating power shafts to both the sun and ring gears. Parts corresponding to those illustrated in FIG. l have the same numbers primed. It will be seen that a second shaft 56 is attached to the sun gear 18', and that a bearing arrangement positions the shaft 25', which is attached to ring gear 13, between housing and shaft 56.

Accordingly it is seen that my invention accomplishes its intended objects, providing a rotary gear pump or motor in which the various gear elements are automatically centered in the clearance area provided for them, to reduce wear and clearance leakage and that cavitation is avoided.

It will be understood that variations and modifications may be effected without departing from the spirit and scope of the novel concepts of this invention.

I claim:

1. A rotary gear motor/pump structure comprising housing means; shaft means rotatably mounted in said housing means extending therefrom; a plurality of gears within said housing means; said shaft means connecting with at least one of said gears; means confining said gears within said housing means in motoring/pumping relationship; fluid bearing recesses on opposite sides of said gears concentric with the centerline thereof; means including said gears defining a plurality of fluid trapping zones and fluid expanding zones; restricted passage means placing said bearing recesses in communication with said plurality of fluid trapping zones on the converging side of a line joining the centers of said gears for delivery of fluid from the latter to the former to provide hydrostatic bearings on opposite sides of said gears acting directly thereagainst to axially center the same.

2. A rotary gear structure according to claim 1, wherein said passage means is divided and feeds hydrostatic bearing recesses on plural gears.

3. A rotary gear structure according to claim 1, wherein said passage means feed the bearing recesses on opposites of the same gear from a common trapping zone.

4. A rotary gear structure according to claim 1, wherein said restricted passage means includes a passage on one side of a gear feeding an adjacent hydrostatic thrust bearing recess and another passage on the opposite side of the gear feeding another recess such that all gears may be axially centered in substantially the mid position of their clearance.

5. A rotary gear motor/pump structure according to claim 1, together with means for supplying fluid to said expanding zones from a source other than said bearing recesses to avoid cavitation.

6. A rotary gear structure according to claim 5 wherein said anticavitation means allows leakage fluid to flow from a drainage area back to said expanding zones.

7. A rotary gear structure according to claim 5, further including additional restricted passage means for bidirectional operation and wherein said expanding zones are supplied with fluid through check valve control means. 

1. A rotary gear motor/pump structure comprising housing means; shaft means rotatably mounted in said housing means extending therefrom; a plurality of gears within said housing means; said shaft means connecting with at least one of said gears; means confining said gears within said housing means in motoring/pumping relationship; fluid bearing recesses on opposite sides of said gears concentric with the centerline thereof; means including said gears defining a plurality of fluid trapping zones and fluid expanding zones; restricted passage means placing said bearing recesses in communication with said plurality of fluid trapping zones on the converging side of a line joining the centers of said gears for delivery of fluid from the latter to the former to provide hydrostatic bearings on opposite sides of said gears acting directly thereagainst to axially center the same.
 2. A rotary gear structure according to claim 1, wherein said passage means is divided and feeds hydrostatic bearing recesses on plural gears.
 3. A rotary gear structure according to claim 1, wherein said passage means feed the bearing recesses on opposites of the same gear from a common trapping zone.
 4. A rotary gear structure according to claim 1, wherein said restricted passage means includes a passage on one side of a gear feeding an adjacent hydrostatic thrust bearing recess and another passage on the opposite side of the gear feeding another recess such that all gears may be axially centered in substantially the mid position of their clearance.
 5. A rotary gear motor/pump structure according to claim 1, together with means for supplying fluid to said expanding zones from a source other than said bearing recesses to avoid caVitation.
 6. A rotary gear structure according to claim 5 wherein said anticavitation means allows leakage fluid to flow from a drainage area back to said expanding zones.
 7. A rotary gear structure according to claim 5, further including additional restricted passage means for bidirectional operation and wherein said expanding zones are supplied with fluid through check valve control means. 